Dimensionally-Optimized Device For The Exchange Of Heat And Method For Optimisation Of The Dimensions Of Devices For The Exchange Of Heat

ABSTRACT

The invention relates to a device for the exchange of heat, in particular, for a motor vehicle, comprising a number of flow tubes for the transport of a fluid, whereby the device has a given depth (T) and, at least in sections, some flow tubes are arranged at a given separation from each other, whereby the depth and the given separation are in a ratio (V) to each other.

The present invention relates to a heat-exchanging device. The device is described in the context of an air-conditioning system, in particular for motor vehicles. It is however pointed out that the device according to the invention can also be used in other air-conditioning systems or refrigeration circuits.

From the prior art, air-conditioning systems or heat-exchanging devices are known which, for cooling, use the refrigerant R 134a. Also known are air-conditioning systems which, instead of said refrigerant, use the refrigerant R 744, that is to say carbon dioxide (CO₂). The advantage of the use of CO₂ over earlier refrigerants is inter alia its better environmental compatibility, since said refrigerant does not lead to an increase in the greenhouse effect.

However, in the prior art, the use of CO₂ for refrigeration circuits is associated with to some extent considerable excess costs over conventional refrigerants, since said refrigerant is under a significantly higher pressure in the device than R 134a. It is thus for example the case that using the same geometry or the same dimensions of the refrigeration circuit as are used for conventional refrigerants results in a very high weight and also high production costs, which lead to the production of the devices becoming uneconomical. It is the object of the invention to adapt individual heat-exchanging devices in terms of their dimensions to the use of CO₂ as refrigerant in such a way as to permit more cost-effective and low-weight production.

Through extensive tests, it was possible to show that a particularly efficient saving of weight and production costs can be obtained by means of modifications to the evaporator. Here, it is also conceivable to accept a moderate reduction in capacity of the evaporator, since this, as has been shown from comprehensive analysis, has the smallest effect on the cooling which can be obtained in the vehicle.

The object is to adapt the devices, by adapting certain dimensions, in such a way as to obtain an improvement of the device, in particular when using the refrigerant CO₂, but in particular not only in terms of its production costs, capacity, weight etc.

A further object is also to improve heat-exchanging devices which use R 134a as refrigerant.

The object is achieved according to the invention in that the evaporator as a component is reduced in terms of its specific refrigerating capacity by an order of magnitude at which the repercussions in the refrigeration circuit in the vehicle cabin are still acceptable. Here, it can be accepted that the capacity level of a refrigeration circuit using a conventional refrigerant (R 134a) is no longer considerably outperformed as before, but is rather at a comparable level. More precisely, the evaporator should be designed so as to be comparable, in terms of its refrigerating capacity, its weight and its production costs, to evaporators with conventional refrigerants.

According to the invention, significant geometric dimensions of the evaporator are optimized so as to obtain the most favorable possible cost/benefit ratio within the context of the entire system.

The object is achieved in detail by means of the subject matter of claim 1. Advantageous embodiments and developments are the subject matter of the subclaims.

The heat-exchanging device according to the invention has a plurality of throughflow tubes for conveying a Fluid, with the device having a predefined depth—also referred to below as the installation depth—and some throughflow tubes are arranged at least in sections at a predefined spacing to one another. In this way, according to the invention, the ratio between the depth and the predefined spacing is less than 7. The depth of the heat-exchanging device results substantially from the depth of the individual throughflow tubes, as will be explained in detail with reference to the figures.

The tube spacing of the individual throughflow tubes is to be understood here, as can be seen more precisely with reference to the figures, as the spacing by which the sides, which face toward one another in each case, of the throughflow tubes are spaced apart from one another. Said tube spacing also determines the height of the fins which are preferably arranged between the tubes. The tube spacing is therefore also referred to below as the fin height.

Here, the spacing is to be understood to mean the shortest geometric spacing between the throughflow tubes. The spacing which is predefined at least in sections is to be understood to mean that the tubes need not strictly have the same spacing from one another along their entire length.

It is also possible for a first group of tubes to have a first spacing to one another and for a second group of tubes to have a second spacing to one another. This will also be explained in detail in connection with the figures.

In a further preferred embodiment, the ratio V is less than 6.5, preferably less than 6.3 and particularly preferably less than 5.9. Through tests and analysis, it was possible to determine that said ratios lead, when using CO₂ as refrigerant, to a particularly favorable cost/benefit relationship, with in particular the specific refrigerating capacity, the air-side and refrigerant-side pressure drop and the production costs and the weight being possible criteria for evaluating the cost/benefit ratio.

In a further preferred embodiment, the throughflow tubes are arranged parallel to one another at least in sections. A substantially constant spacing between the individual throughflow tubes can be ensured in this way.

The throughflow tubes are preferably parallel to one another substantially along their entire length and in this way have a constant predefined first spacing to one another along substantially their entire length.

In a further preferred embodiment, the throughflow tubes have a flat-tube-shaped cross section. A flat-tube-shaped cross section is to be understood as a cross section in which one side by far exceeds a further side in terms of its length, such as for example an elongated rectangle, an elongated rectangle with rounded corners or an ellipse in which the first diameter is considerably greater than the second diameter.

In a further preferred embodiment, the fluid is a refrigerant and is preferably R 744 (CO₂).

In a further preferred embodiment, a first plurality of throughflow tubes have a first predefined spacing to one another at least in sections, and a second plurality of throughflow devices have a second, substantially predefined spacing to one another, with the ratio V between the depth and at least one of the predefined spacings being less than 7.

This means that individual throughflow devices have a different spacing from one another than other throughflow devices. Here, the individual spacings can also vary within throughflow devices which are substantially parallel to one another. In addition, both the first predefined spacing and the second predefined spacing can be dimensioned such that the ratio between the depth and both predefined spacings is in each case less than 7.

It is preferable, in a heat-exchanging device in which the fluid is conveyed in a certain direction in a first plurality of throughflow devices and in a further plurality of throughflow devices, in which the fluid is conveyed in another direction, for in each case different predefined tube spacings to be selected. In this way, it is possible under some circumstances to obtain a more cost-effective design in terms of the obtained heat transfer capacity.

In a further preferred embodiment, a first plurality of throughflow devices is laterally offset with respect to a second plurality of throughflow devices. Here, the individual predefined spacings of the first and of the second plurality can be selected to be the same or different. The predefined spacings can also vary within the same plurality of throughflow devices.

In a further preferred embodiment, cooling fins are arranged between the throughflow tubes. Said cooling fins serve to improve the exchange of heat with the surrounding air. Here, as mentioned, the height of said cooling fins is substantially determined by the predefined spacing of the respective throughflow tubes which bear against them.

In one preferred embodiment, the wall thickness of the individual cooling fins is between 0.04 and 0.2 mm, preferably between 0.05 and 0.12 mm and particularly preferably between 0.06 and 0.1 mm. The fin density is between 40 and 90 fins/dm, preferably between 50 and 80 fins/dm and particularly preferably between 60 and 70 fins/dm.

In a further preferred embodiment, the device has a depth of between 10 mm and 60 mm, preferably between 20 mm and 50 mm, and particularly preferably between 25 and 45 mm. Said different depths are determined in particular by the intended application, that is to say for example by whether the device is to be used in a compact vehicle, a medium-sized vehicle or a large vehicle.

In a further preferred embodiment, the predefined spacing between the throughflow tubes is between 4 mm and 12 mm, preferably between 4.5 mm and 10 mm. Said spacings are also determined in particular by the respective applications.

In a further preferred embodiment, a depth of between 30 mm and 50 mm, preferably a depth of between 35 mm and 45 mm is assigned a predefined spacing of between 5 mm and 12 mm, preferably between 5.5 mm and 10 mm. Said embodiment concerns more largely-dimensioned heat-exchanging devices which can be used in particular, but not exclusively, in air-conditioning systems in medium-sized vehicles or large vehicles. Here, however, the selected dimensions substantially ensure that the ratio remains lower than 7.

In a further preferred embodiment, a depth of between 15 mm and 40 mm, preferably a depth of between 20 mm and 35 mm is assigned a predefined spacing of between 3 mm and 10 mm, preferably between 4 mm and 8.5 mm. Said dimensions or measurements are used in particular in air-conditioning systems of compact vehicles and medium-sized vehicles.

Said dimensions, too, should substantially ensure a ratio of less than 7. Here, however, a ratio of substantially 7 is also to be understood as a ratio which slightly exceeds the value 7.

In a further preferred embodiment, the throughflow tubes have a width of between 1 mm and 3 mm, preferably between 1.5 mm and 2 mm and particularly preferably between 1.7 mm and 1.9 mm. The wall thickness of the throughflow tubes is between 0.1 mm and 0.6 mm, preferably between 0.2 mm and 0.4 mm and particularly preferably in the region of approximately 0.3 mm. With said dimensions, it is possible to obtain a particularly advantageous exchange of heat with the ambient air.

The device according to the invention is preferably an evaporator which is a component of a refrigeration circuit of a motor vehicle air-conditioning system.

The invention is also aimed at an air-conditioning system, in particular for a motor vehicle, which has at least one heat-exchanging device according to the invention.

The invention is also aimed at a method for dimensioning heat-exchanging devices, in which method a first dimension of the device is specified in a first step, a second dimension of the device is specified in a further step, at least two first target parameters of the device are determined in a further step, at least one dimension is varied in a further step, two second target parameters of the device with the varied dimension are in turn determined from the varied dimension, and finally the more favorable target parameters are selected by comparing the first and second target parameters.

The first and second dimensions are preferably selected from a group of dimensions which contains the depth, the fin height of the cooling fins and the spacing of the throughflow tubes and the like.

Dimensions can however also be understood to mean variables such as the fin density per dm and the like.

Target parameters are preferably selected from a group of parameters which contains the installation space depth, the refrigerating capacity, the air-side pressure drop, the weight and the production costs. As mentioned in the introduction, said factors ultimately determine the benefit or value of the heat-exchanging device for the different refrigerants, in the present case for R 134a and R 744 (CO₂). Using the method according to the invention, the significant dimensions of the heat-exchanging device can be varied and thereby in each case the related stated output variables determined in order to thereby arrive at a device which is dimensioned so as to provide satisfactory, sufficient refrigerating capacity with acceptable weight, with acceptable production expenditure and acceptable costs.

In said method, it is to be taken into consideration that even small changes to one or the other dimension can lead to drastic changes in an output variable or a target parameter.

The target parameters are preferably determined multiple times in particular for different dimensions, and, from this plurality of determined sets of target parameters which are determined in this way, the most favorable sets of parameters are determined. With said multiple determination of the target parameters, it is possible to provide very accurate analysis of the capacities or target parameters of the heat-exchanging device which can be expected. In the determination of the most favorable target parameter sets, the individual target parameters are preferably weighted according to predefined criteria. It is thus for example possible, where the device is to be used in a large vehicle, for the weight and production cost target parameters to be weighted lower than in the case of application in a compact vehicle.

Further advantages and embodiments of the device according to the invention and of the method according to the invention can be gathered from the figures, in which:

FIG. 1 shows a plan view of a detail of the device according to the invention;

FIG. 2 shows a side view of the device according to the invention from FIG. 1;

FIG. 3 is a schematic illustration of a further embodiment;

FIG. 4 is a schematic illustration of a further embodiment;

FIG. 5 is a schematic illustration of a further embodiment;

FIG. 6 is a schematic illustration for clarifying the tube spacings;

FIG. 7 shows a diagram for clarifying the cooling which is obtained;

FIG. 8 is an illustration for analyzing the individual components;

FIG. 9 a is a graphic illustration of the ratio between refrigerating capacity and weight of the device according to the invention;

FIG. 9 b is an illustration of the air-side pressure drop;

FIG. 10 is an illustration of the capacity as a function of the installation depth;

FIG. 11 is an illustration of the ratio of capacity to weight as a function of the installation depth;

FIG. 12 is an illustration of the capacity in relation to costs as a function of the installation depth;

FIG. 13 is an illustration of the capacity as a function of the installation depth to fin height;

FIG. 14 is an illustration of the capacity in relation to the weight as a function of the installation depth in relation to the fin height; and

FIG. 15 is an illustration of the capacity in relation to the costs as a function of the ratio of installation depth to fin height.

FIG. 1 shows a plan view of a detail of the heat-exchanging device 1 according to the invention. Said heat-exchanging device 1 has a plurality of first throughflow tubes 3 and a second plurality of second throughflow tubes 5. In a preferred embodiment, the refrigerant flows through the plurality of first throughflow tubes 3 in one direction, for example out of the plane of the page, and in an opposite direction, that is to say into the plane of the page, in the second plurality of throughflow tubes 5.

The reference sign 7 denotes a chamber of the throughflow tube. The throughflow tubes are preferably divided into a plurality of chambers or ducts.

Here, the first throughflow tubes 3 and second throughflow tubes 5 are separated from one another by an intermediate space 8. Said intermediate space 8 serves for heat insulation, since the refrigerant in the throughflow tubes 3 and 5 can have a different temperature, and heat transfer should not take place. Instead of the intermediate space, it is however also possible for the throughflow tubes to be arranged continuously along the depth T, that is to say for only one plurality of flat tubes to be provided. In this case, one chamber or one duct 7 is preferably designed to be dead-ended, that is to say no refrigerant flows in said duct.

The reference symbol 4 relates to fins which are arranged between the throughflow tubes 3 and 5 and are shown here in a plan view from above. The dimension H_(Ri) denotes the fin height and is determined substantially by the spacing of the individual throughflow tubes 3 and 5, more precisely by the spacing of the sides, which face toward one another in each case, of the respective throughflow tubes 3 and 5.

The reference symbol T denotes the installation depth which, as mentioned above, constitutes a significant geometric variable of the device. The fins 4 extend substantially along the entire depth T and are preferably also not interrupted by intermediate spaces. The ratio V mentioned above determines the ratio of installation depth T to the fin height H_(Ri).

FIG. 2 shows a side view of the illustration, of which a detail is shown in FIG. 1, of the heat-exchanging device. Here, b denotes the tube width of the individual throughflow tubes. In the case of a heat-exchanging device which utilizes R 134a as refrigerant, the width of the tubes is between 2 and 4 mm, preferably between 2.5 and 3 mm.

In a heat-exchanging device which utilizes CO₂ as refrigerant, the width of the tubes is preferably—as mentioned above—in a range from 1.2 to 2 mm. The device has an overall width of between 120 and 400 mm, preferably between 215 and 350 mm and particularly preferably between 250 and 315 mm. A further advantageous width is between 120 and 315 mm. The height of the device according to the invention is between 140 and 300 mm, preferably between 200 and 300 mm, particularly preferably between 220 and 250 mm. A further advantageous height is between 140 and 270 mm. In one preferred embodiment, the device is produced substantially from aluminum or an aluminum-containing material.

The reference symbol A denotes the so-called transverse pitch, that is to say the spacing of the respective geometric centers of the individual throughflow devices to one another. Said transverse pitch A results in the fin height H_(Ri) when the respective tube width b is also taken into consideration, that is to say the fin height and the transverse pitch are directly related. The transverse pitch can be incorporated as a measure for the fin height if, on account of the cross section of the throughflow tubes 3, 5, there is no geometrically unequivocal and constant value for the fin height or for the spacing of the throughflow tubes, for example if the spacing of the throughflow tubes in FIG. 2 varies in a direction perpendicular to the plane of the page, which is for example possible in the case of a circular profile of the throughflow tubes. In this case, the ratio according to the invention of the depth and spacing of the tubes is to be replaced by the ratio of depth and transverse pitch.

FIG. 3 schematically illustrates a further embodiment of the device according to the invention. Here, the reference symbols 3 and 5 relate in each case to plan views of the individual throughflow tubes. In contrast to the embodiment shown in FIG. 1, here, the throughflow tubes 3 and the throughflow tubes 5 are laterally offset with respect to one another. This means that the spacing between the throughflow tubes can be determined separately for the throughflow tubes 3 and for the throughflow tubes 5. In the exemplary embodiments shown in FIG. 3, the spacing H_(Ri) the throughflow tubes 3 is identical to the spacing H_(Ri) the throughflow tubes 5.

FIG. 4 schematically illustrates a further embodiment of the device according to the invention. In this case, the throughflow tubes 3 have a greater spacing H_(ri) from one another than the throughflow tubes 5 which have a spacing of H_(Ri2) from one another. Here, preferably at least one of the two spacings H_(Ri1) or H_(Ri2), in this case at least the spacing H_(Ri1), is selected such that the ratio of the depth T and the spacing H_(Ri1) is less than 7. It is however also possible to select both spacings such that the corresponding ratio is less than 7.

FIG. 5 illustrates a further embodiment of the device according to the invention. In this embodiment, the spacings between the individual throughflow tubes vary only within the throughflow tubes 3. It is however also possible for the spacings to vary only within the tubes 5, or else both within the throughflow tubes 3 and the throughflow tubes 5. In this embodiment, too, it must be ensured that at least one of the spacings H_(Ri) meets the criterion that the ratio of the depth and said spacing is less than 7.

It would also be possible to provide further different spacings or a plurality of different spacings between the individual tubes, such as for example spacings H_(Ri1), H_(Ri2), H_(Ri3), etc. In any case, the above ratio, which is less than 7, would have to be maintained for one of the spacings H_(Ri).

FIG. 6 is a schematic illustration for clarifying the definition of the spacing H_(Ri). While the throughflow tubes in FIGS. 3 to 5 in each case have rectilinear longitudinal sides, which simultaneously directly determine the spacing, the throughflow tubes in the embodiment shown in FIG. 6 have an elliptical cross section. In this case, the spacing between the throughflow tubes is defined as the spacing of the two tangents T which are each placed against the throughflow tubes 3.

It is however also possible, as indicated above, to define the tube spacing not by means of the spacing of the sides which face toward one another, but rather by means of the spacing of the respective geometric center line of the individual throughflow tubes, as was referred to above as the transverse pitch. This lends itself to use, as stated, primarily when throughflow tubes have geometries which deviate from the geometries shown here, such as for example concave or convex shapes.

The diagram shown in FIG. 7 shows the simulation of a cooling curve for a large vehicle. Here, comparable cooling curves have been illustrated for the coolant R 134a, here by curves 11 and 12, and for R 744, here illustrated by the curves 14 and 15, in each case in the idle operating point.

The upper curves 12 and 14 show the temperature profile in the vehicle interior, and the lower curves 11 and 15 show the temperature generation at the evaporator itself.

For the simulation, it was also assumed that the R 744 evaporator has an installation depth which is 25 mm smaller, specifically an installation depth of 40 mm, whereas the R 134a evaporator has an installation depth of 65 mm.

Plotted on the ordinate is the time in minutes, and on the abscissa the temperature in degrees Celsius. The simulation is divided into several time sections I to IV, with section I relating to driving in 3^(rd) gear at 32 km/h, section II relating to driving in 4^(th) gear at 64 km/h, section III relating to idle, and section IV relating to driving in 2^(nd) gear at 64 km/h.

It can be seen that even in 3^(rd) gear (I) the R 744 evaporator provides more efficient cooling than the R 134a evaporator. In regions II to IV, the respective evaporators obtain in each case substantially the same values.

FIG. 8 illustrates a capacity comparison of different evaporator designs at a typical operating point. Here, said operating point is defined so as to permit comparisons which are independent of the refrigeration circuit.

It is pointed out that the method and the obtained results which are described in the following can be incorporated equally for improving both R 134a evaporators and also R 744 (CO₂) evaporators.

In the diagram illustrated here, the assumptions are made of an air mass flow rate GLV of 8 kg/min, an air inlet temperature tLVe=40° C., and a relative humidity φLVE of 40%.

In the diagram, the rhombuses show the values determined for the refrigerant R 744 (CO₂) and the ellipses show the values determined for the refrigerant R 134a.

The fin density is 70 fins/dm for the evaporator with the refrigerant R 744, and 60 fins/dm for the evaporator with the refrigerant R 134a.

Plotted on the ordinate is the installation depth in mm, and on the abscissa the total capacity in kW. The plotted value pairs or points 31 to 39 are functions of the temperature T, the fin height H_(Ri), the fin density z_(ri) and the so-called transverse pitch s_(q). The transverse pitch denotes the spacing of the respective centers of the individual throughflow tubes from one another. Here, the individual value pairs or points 31 to 39 span a field which covers the capacity level in refrigeration circuits of different vehicle classes. Here, the upper curve 22 is the large vehicle or van segment, and the lower limit curve 23 shows the capacity requirement of compact vehicles.

The values for the refrigerant R 744 are plotted for installation depths smaller than 40 mm, that is to say for the measurement points 31 to 35. The values for the refrigerant R 134a are plotted for the installation depth region Σ 40 mm. As mentioned above, a uniform fin density of 70 fins/dm was selected for the measurement points 31 to 35, while a uniform fin density of 60 fins/dm was selected for points 36 to 39.

In the case of the measurement points 31 and 32, a relatively small transverse pitch was used, and in the case of the measurement points 33 to 35, a relatively high transverse pitch was used. The relatively small transverse pitch results in a likewise small fin height, which is indicated by the line 28. The relatively high transverse pitch likewise results in a relatively high fin height, which is shown by the line 27.

For the measurement points 36 and 37, a relatively small transverse pitch was selected, which leads to a relatively small fin height H_(Ri), as shown by the line 25. For the measurement points 38 and 39, a relatively high transverse pitch was selected, which leads to a relatively high fin height, as shown by the line 26.

It can be seen from the diagram that the installation depth for R 744 is considerably reduced for a constant capacity level, which is plotted on the abscissa. This means that the assignment of the installation depth T to fin height H_(Ri) or the ratio is moved.

While, in the case of R 134a, a depth of 65 mm is assigned to a fin height of 7 to 10 mm and a depth of 40 mm is assigned to a fin height of 4 to 6 mm, when using the refrigerant R 744, a depth of 40 mm is assigned to a fin height of 7 to 10 mm and a depth of 27 mm is assigned to a fin height of 5 to 8 mm. In earlier designs, for the refrigerant R 744, the assignment or dimensioning of R 134a was assumed. This led to considerably higher capacity values than R 134a but also to excess weight and to excess costs, which is caused inter alia by the considerably higher pressures required with R 744. Said considerably higher capacity values are plotted by way of example by the points 41 and 42. The capacities at points 41 and 42 are around over 15′ above the maximum required capacities.

It was thereby possible to show that, counter to the expectations among experts, cost-reducing and weight-reducing modifications in dimensioning are possible without thereby having to simultaneously accept reductions in cooling capacity.

The considerably higher potential in R 744 is based on the fact that, on account of the high specific delivery rate of the R 744 compressor in the R 744 circuit, a faster pressure reduction is obtained in the low-pressure part. This leads to improved dynamics and, at the evaporator, to a higher driving pressure gradient between the air and the refrigerant.

The refrigerant-side pressure drop in the evaporator is of a comparable order of magnitude, with 1 bar pressure drop with R 134a causing approximately 9K temperature variation, and with R 744, only 1K. This leads, on average, to a considerably higher driving pressure gradient between the air and the refrigerant over the flow length in the evaporator (the R 744 evaporator provides, on average, a considerably colder surface temperature).

As mentioned in the introduction, it is sought to provide a cost/benefit optimum as a function of the variables of installation space depth, refrigerating capacity, air-side pressure drop, weight and costs. Here, as mentioned in the introduction, the variables are the depth T, the fin height H_(Ri) and the tube spacing, and variables derived from said variables, like the transverse pitch.

According to previous considerations and tests, an installation depth of 65 mm for the present capacity level is rather too large; according to estimations, a 55 mm deep design which reaches the level of the 65 mm depth would be more favorable. However, an embodiment of said type possibly leads to higher costs and to a less favorable air-side pressure drop. For the refrigerant R 134a, a depth of 40 mm has proven to be particularly favorable in terms of the capacity; in this case, however, disadvantages with regard to costs and the air-side pressure drop are to be expected. Said considerations show the extremely complex interrelationships between different aspects in the assessment and evaluation of the evaporators to be produced.

In the case of an evaporator which utilizes the refrigerant R 744, an installation space depth of between 25 and 45 mm has proven to be particularly suitable.

The diagram shown in FIG. 9 displays some of the advantages of the invention. Here, the sub-diagram denoted by FIG. 9 a illustrates the weight of the evaporator in relation to the obtainable refrigerating capacity. The physical boundary conditions such as for example the air mass flow rate GLV are identical to the conditions which formed the basis in the description of FIG. 8. The same evaporator dimensions were likewise selected.

As shown by the measurement points 44 and 45, which relate to compact vehicles and medium-sized vehicles, it is possible by adapting the geometric dimensions to obtain comparable refrigeration capacities, with the measurement point 44 having been determined for the refrigerant R 744 and the measurement point 45 having been determined for the refrigerant R 134a. In the case of the measurement point 45, a medium installation depth and a fin density of 60 fins/dm was used. In the case of the measurement point 44, a smaller depth, a smaller transverse pitch and a higher fin density were used than in the case of point 45.

The two measurement points 46 and 47, which relate to devices for large vehicles, also show, for the same refrigerating capacity, a considerably reduced weight of the R 744 evaporator. For the measurement point 46, a relatively high installation depth T, a predefined fin density and a relatively high transverse pitch Sq were selected. In the case of the measurement point 47 for the R 744 evaporator, a smaller depth T than at point 46, the same fin density as at point 46 and a correspondingly identical transverse pitch were selected. The result is therefore a considerable weight reduction, as a result of the smaller installation depth with otherwise identical transverse pitch, and even weight advantages over the respective identical-capacity R 134a evaporator. The smaller installation depth also results in only relatively low material costs, and therefore a cost reduction.

In addition, it is possible for evaporators for large vehicles to obtain an installation depth reduction from 65 to 40 mm, and in the case of compact vehicles, from 40 to 25 mm. This brings with it the additional advantage that less installation space is required in the motor vehicle.

As the diagram in FIG. 9 b shows, the air-side pressure drop, which is illustrated on the abscissa, can also be reduced. The blocks 51 to 53 relate to the refrigerant R 134a, the blocks 54 to 55 to the refrigerant R 744. It can be seen that, when using R 744, a considerable reduction in the air-side pressure drop by approximately 50% is also obtained. This leads to an increased air quantity for the air-conditioning of the vehicle, to a reduced power consumption in the fan and also offers the potential for reducing the noise level in the air-conditioning unit.

In FIG. 10, the capacity values of individual evaporators are plotted against the installation depth on the ordinate. Here, both for the CO₂ and for the R 134a evaporators, the evaporators with the same fin height are situated in each case on one line. The reference symbol 63 denotes the line which is associated with a large fin height, which is referred to below as the first fin height, the reference symbol 62 denotes the line which is associated with a second, smaller fin height (referred to below as the second fin height), and the reference symbol 61 denotes the line which is associated with a fin height (referred to below as the third fin height) which is smaller still than the second fin height.

As can be seen from FIG. 10, the individual lines 61 to 63 have relatively similar gradients, which suggests a proportional dependency of capacity and installation depth for an otherwise identical design or fin height. It can also be seen that evaporators with smaller fin heights but otherwise identical dimensions have higher capacities on account of the enlargement of the heat-transmitting surface.

The hatched regions 60 and 70 delimit the required or expedient capacity values. The capacity limits have been determined inter alia by means of the simulation of vehicle cabin cooling. While, in the upper region 60, a further capacity increase no longer brings with it any advantages, the cabin cooling below the lower limit in the region 70 is no longer acceptable. The reference symbols 65 to 68 show measurement values which lie within the required capacity region. Said reference symbols 65 to 68 denote devices of different design.

The reference symbol 67 relates to an R 134a evaporator with a large installation depth and the first above-stated fin height. The reference symbol 65 refers to an R 134a evaporator with the third above-stated fin height and a relatively small installation depth.

The reference symbol 66 relates to an R 134a evaporator with the second fin height and a medium installation depth.

The reference symbol 68 refers to an R 134a evaporator with the first fin height and a medium installation depth. The reference symbols 71 to 74 show the measurement values of those evaporators which no longer lie in the tolerable range 75 between the regions 60 and 70. Here, the reference symbol 71 denotes an evaporator with a small installation depth and the first fin height, the reference symbol 72 denotes an R 744 evaporator with the third fin height and a very small installation depth, the reference symbol 73 denotes a CO₂ evaporator with a small installation depth and the third fin height, and the reference symbol 74 denotes a CO₂ evaporator with a large installation depth and the first fin height.

It can therefore be seen that the CO₂ evaporators have significantly higher capacity values at the given installation depths and fin heights than the R 134a evaporators. As can likewise be seen from FIG. 10, a CO₂ evaporator which has a small installation depth for example with the second fin height, as shown by the line 76, could be of interest for the application. The ellipses 140, 141 indicate regions in which favorable dimensions are situated.

FIG. 11 illustrates the relationship of capacity to weight as a function of the installation depth. Here, the related variables capacity/weight have again been weighted relative to one another in order to give consideration to the different significance of the individual variables. In a further preferred variant of the method, the capacity and the costs are considered as variables of equal significance, while the weight and the fin height are of secondary importance.

Here, in the diagrams shown in FIGS. 11 to 15, the weightings have been applied in such a way that the capacity is weighted in the ratio 50:50 in relation to the costs, the capacity is weighted in the ratio 80:20 in relation to the weight, and the installation depth is weighted in the ratio 70:30 in relation to the fin height. The triangles relate in each case to CO₂ evaporators, and the circles to R 134a evaporators. Since the ratio capacity/weight is plotted on the abscissa, higher values, that is to say a ratio displaced more in the direction of capacity, are to be considered more favorable.

It can thereby be seen that, for the R 134a evaporators, the evaporator denoted by the reference symbol 81, with a medium installation depth and an above-stated second fin height, and the evaporator denoted by the reference symbol 83, with a small installation depth and the third fin height, prove to be particularly favorable.

Although the evaporator denoted by the reference symbol 84, with the first fin height, would also lie favorably in relation to the ratio of capacity to weight, its absolute capacity would however no longer be acceptable even for the cooling of compact vehicles. Said evaporator type could conceivably be used for example in a rear system. An evaporator with the same fin height in the region of the second fin height could likewise for example be considered as a further alternative for compact and/or medium-sized vehicles.

For the refrigerant CO₂, therefore, the evaporators denoted by the reference symbols 86 and 87, with a small depth with a large fin height, and the evaporator denoted by the reference symbol 88, with a small depth and the third fin height, are most favorable. Finally, the evaporator denoted by the reference symbol 89, with a relatively small depth, also lies in a relatively favorable position, but is borderline in terms of its capacity.

The evaporator 91 is less favorable in a direct comparison with the first fin height. In addition, said evaporator is already situated above the presently required upper capacity limit.

The evaporator denoted by the reference symbol 92 has an unfavorable ratio of capacity to weight on account of the high packing density of tubes and fins, with an otherwise excessively low capacity.

The reference symbols 95 and 96 relate to trend lines which have been specified on the basis of the measured values. On the basis of said trend lines, it is possible to determine or estimate those dimensions of the evaporator with which favorable designs, such as here a favorable capacity/weight ratio, are to be expected.

In detail, the trend line 95 relates to CO₂ evaporators and the trend line 96 relates to R 134a evaporators.

FIG. 12 illustrates the ratio of capacity to production costs as a function of the installation depth. Here, in the case of the ratio of capacity to costs, the abovementioned ratio or the above-specified weighting has again been incorporated.

It can be seen that, in the case of the R 134a evaporators, which are denoted by circles, the evaporator 101 with a medium installation depth and a first fin height has the best capacity/cost ratio. However, said evaporator has a low output capacity and is therefore not incorporated in generating the trend line 115.

The trend line 115 for the R 134a evaporators and the trend line 116 for the CO₂ evaporators in each case indicate, as above, the geometries with which particularly favorable results for the evaporators are to be expected. Although the evaporator denoted by the reference symbol 102 with a third fin height performs considerably less favorably, here, the advantage of the small installation depth in relation to the evaporators denoted by the reference symbols 104-106 is to be taken into consideration.

In the consideration of the CO₂ evaporators which are denoted by the triangles, the good capacity/cost ratio of the evaporators denoted by the reference symbols 107 and 108, with the first fin height or a fin height which is slightly below said first fin height, is to be maintained, but also that of the evaporator denoted by the reference symbol 110 with the first fin height.

The evaporator denoted by the reference symbol 111 with the third fin height lies, according to expectation, in a slightly less favorable position on account of the high packing density, which has an adverse effect on the cost aspect. An evaporator with the second fin height would logically lie between those with the third and the first fin heights, and would by all means represent an alternative of interest.

Finally, less favorable still is said ratio in the evaporator denoted by the reference symbol 112 with a large installation depth with the third fin height, and in the evaporator denoted by the reference symbol 113 with a small installation depth.

In the former, the high costs are prevalent on account of the small fin height (or the high packing density), while in the latter, the low capacity at still moderate costs is prevalent. The evaporator denoted by the reference symbol 114, which corresponds to the evaporator denoted by the reference symbol 93 in FIG. 11, again remains unconsidered for the above-mentioned reasons.

The general result is a lower level of the CO₂ evaporators than the R 134a evaporators. Here, a certain cost deficit is also to be recognized, which can however be accounted for by a more stable design for strength and safety reasons (considerably higher operating pressures when using CO₂ as refrigerant) and therefore higher weight.

The illustrations shown in FIGS. 13 to 15 correlate to the first illustrations shown in FIGS. 10 to 12. However, in the illustrations shown in FIGS. 13 to 15, the variable “installation depth” plotted on the ordinate or abscissa has been replaced by the weighted ratio V′ of installation depth and the sum of fin height+10 mm.

From the capacity of the individual evaporators over the weighted ratio V′ of installation depth to fin height, as illustrated in FIG. 13, it can be seen that all evaporators operated with the same refrigerant (R 134a and CO₂) are now situated on a constant line substantially independently of their fin height. This explains the selected weighting of the installation depth and fin height relative to one another, which takes the form of the summing of 10 mm to the fin height. Also to be seen again is the capacity advantage of the CO₂ evaporators over the R 134a evaporators of the same installation depths. The individual values are again, as above, measured values or values determined by simulation which have been confirmed by measurements.

The same statements as above also result, in principle, in the illustration shown in FIG. 14 of the capacity related to the weight over the installation depth related to the weighted fin height in relation to the absolute installation depth (cf. FIG. 11). It has moreover been shown that the capacity related to the weight of the tested R 744 evaporators is a maximum between V′=1.3 and V′=2.8, and appears to drop outside said range. Better values are shown for evaporators above V′=1.5, and values which are better still for evaporators above V′=1.85. The evaporators with the highest capacity related to the weight have a weighted ratio V′ of 2.2 and 2.4. The trend line, in contrast, shows a maximum at approximately V′=2.1.

The same applies in turn for the illustration, shown in FIG. 15, of the capacity/cost ratio plotted over the related installation depth in a comparison (cf. FIGS. 12 and 15). Here, too, the preferences do not change. The capacity, related to the costs, of the R 744 evaporators exceeds that of the R 134a evaporators when the weighted ratio V′ is less than approximately 2.6.

It can be seen that, with the method according to the invention, in which, as a function of the predefined dimensions or parameters, that is to say the installation depth and the fin height, different target parameters such as the costs, the power and the weight can be determined and weighted relative to one another, in particular by means of different weighting, which variants ultimately represent the most favorable embodiments. In this way, it is possible with the method according to the invention, using different application methods which partially also incorporate weighting, for the most favorable dimensions for the R 134a and the CO₂ evaporators to be worked out in a particularly efficient way. It is possible in this way for the ideal dimensioning for the individual evaporators to be selected taking into consideration the criteria such as weight, capacity etc.

Specially-developed programs are preferably used for the method, which programs allow the user to specify desired criteria, to specify the desired target parameters, in order to thereby satisfy the requirements for example of the air-conditioning of a motor vehicle. It is necessary in generating a program of said type to introduce or incorporate the experience gained in each case through measurement and/or complex thermodynamic considerations.

The invention is therefore also aimed at an item of software which allows the method according to the invention to be carried out on a computer.

For the CO₂ evaporators, particularly favorable installation depths were determined in the range from 20 to 45 mm with a fin height of 4.0 to 10.0 mm.

Proven to be particularly advantageous were installation depths between 35 and 45 mm with fin heights of 5.5 to 10 mm, in particular for use in large vehicles, and installation depths from 20 to 35 mm with fin heights from 4 to 8.5 mm, in particular for use in compact and medium-sized vehicles. 

1. A heat-exchanging device for an air-conditioning system, in particular for a motor vehicle, having a plurality of throughflow tubes for conveying a fluid, with the device having a predefined depth T and a predefined number of throughflow tubes are arranged at least in sections with a predefined spacing to one another, wherein the ratio V between the depth and the predefined spacing is less than 7 and/or the weighted ratio V′ between the depth and the sum of the predefined spacing and 10 mm is greater than 1.3 and less than 2.8.
 2. The device as claimed in claim 1, wherein the predefined spacing is less than or equal to 9 mm, preferably less than or equal to 8 mm and preferably less than or equal to 6 mm.
 3. The device as claimed in claim 1, wherein the ratio V is less than 6.8, preferably less than 6.6 and preferably less than 6.3.
 4. The device as claimed in claim 1, wherein the ratio V is less than 6.1, preferably less than 5.9 and preferably less than 5.1.
 5. The device as claimed in claim 1, wherein the weighted ratio V′ is at least 1.5, preferably at least 1.85 and preferably at least 2.2.
 6. The device as claimed in claim 1, wherein the weighted ratio V′ is at most 2.6, preferably at most 2.4 and preferably at most 2.25.
 7. The device as claimed in claim 1, wherein the throughflow tubes are arranged parallel to one another at least in sections.
 8. The device as claimed in claim 1, wherein the throughflow tubes have a substantially constant predefined first spacing to one another.
 9. The device as claimed in claim 1, wherein the throughflow tubes have a flat-tube-shaped cross section.
 10. The device as claimed in claim 1, wherein the throughflow tubes are formed in one piece, in particular from a single sheet metal strip or extruded profile.
 11. The device as claimed in claim 1, wherein the throughflow tubes and preferably the device have a rupture pressure of over 90 bar.
 12. The device as claimed in claim 1, wherein the fluid is a refrigerant and is preferably R 744 (CO₂).
 13. The device as claimed in claim 1, wherein a first plurality of throughflow tubes have a first predefined spacing to one another at least in sections, a second plurality of throughflow tubes have a second predefined spacing to one another, and the ratio V between the depth and at least one of the predefined spacings is less than
 7. 14. The device as claimed in claim 1, wherein a first plurality of throughflow tubes is laterally offset with respect to a second plurality of throughflow tubes.
 15. The device as claimed in claim 1, wherein cooling fins are arranged between the throughflow tubes.
 16. The device as claimed in claim 1, wherein said device has a depth of between 10 mm and 60 mm, preferably between 20 mm and 50 mm, and particularly preferably between 25 mm and 45 mm.
 17. The device as claimed in claim 1, wherein said device the predefined spacing is between 4 mm and 12 mm, preferably between 4.5 mm and 10 mm.
 18. The device as claimed in claim 1, wherein a depth of between 30 mm and 50 mm, preferably a depth of between 35 mm and 45 mm is assigned a predefined spacing of between 5 mm and 12 mm, preferably between 5.5 mm and 10 mm.
 19. The device as claimed in claim 1, wherein a depth of between 20 mm and 35 mm, preferably a depth of between 25 mm and 30 mm is assigned a predefined spacing of between 3 mm and 10 mm, preferably between 4 mm and 8 mm.
 20. The device as claimed in claim 1, wherein the throughflow tubes have a width of between 1 mm and 3 mm, preferably between 1.3 mm and 2 mm and particularly preferably from approximately 1.4 mm to 1.9 mm.
 21. The device as claimed in claim 1, wherein the throughflow tubes have a wall thickness of between 0.1 mm and 0.6 mm, preferably of between 0.2 mm and 0.4 mm and particularly preferably of approximately 0.25 mm to 0.3 mm.
 22. The device as claimed in claim 1, wherein the device is an evaporator.
 23. An air-conditioning system, in particular for a motor vehicle, wherein said air-conditioning system has at least one heat-exchanging device as claimed in claim
 1. 24. A method for dimensioning heat-exchanging devices, having the following steps: specifying a first dimension of the device; specifying a second dimension of the device; determining at least two first target parameters of the device; varying at least one dimension; determining at least two second target parameters of the device with the varied dimension; determining the more favorable target parameters by comparing the first and second target parameters.
 25. The method as claimed in claim 24, wherein the dimensions are selected from a group of dimensions which contains the depth, the fin height and the spacing of the throughflow tubes.
 26. The method as claimed in claim 24, wherein the parameters are selected from a group of parameters which contains the installation space depth, the refrigerating capacity, the volume flow rate, the air-side pressure drop, the weight and the production costs.
 27. The method as claimed in claim 24, wherein the target parameters are determined multiple times and, from the sets of target parameters which are determined in this way, the most favorable sets of parameters are determined.
 28. The method as claimed in claim 24, wherein in the determination of the most favorable target parameter sets, the individual target parameters are weighted according to predefined criteria. 